The global market for efficient power generation equipment has been expanding in recent years and is anticipated to continue to expand in the future. The gas turbine combined cycle power plant is a preferred choice for this type of equipment due to relatively low plant investment costs and continuously improving operating efficiency of the gas turbine-based combined cycle which minimizes electricity production costs.
By way of background and with reference to the schematic illustration of FIG. 1, a typical combined cycle gas turbine includes, in serial-flow relationship, an air intake or inlet, a compressor, a combustor, a turbine, a heat recovery steam generator (HRSG) and its associated steam turbine. Thus, air enters the axial flow compressor at 10 at ambient conditions. Ambient conditions vary from one location to another and day to day. Therefore, for comparative purposes standard conditions are used by the gas turbine industry. Those standard conditions are 59.degree. F./15.degree. C., 14.696psia/1.013 bar and 60% relative humidity. The standard conditions were established by the International Standards Organization ("ISO") and are generally referred to as ISO conditions.
The compressed air enters the combustion system at 12 where fuel is injected and combustion occurs. The combustion mixture leaves the combustion system and enters the turbine at 14. In the turbine section, energy of the hot gases is converted into work. This conversion takes place in two steps. The hot gases are expanded and the portion of the thermo-energy is converted into kinetic energy in the nozzle section of the turbine. Then, in the bucket section of the turbine a portion of the kinetic energy is transferred to the rotating buckets and converted to work. A portion of the work developed by the turbine is used to drive the compressor whereas the remainder is available for generating power. The exhaust gas leaves the turbine at 16 and flows to the HRSG.
The Brayton cycle is the thermodynamic cycle upon which all gas turbines operate. Every Brayton cycle can be characterized by pressure ratio and firing temperature. The pressure ratio of the cycle is the compressor discharge pressure at 12 divided by the compressor inlet pressure at 10. The General Electric Co. (GE), and we, define the firing temperature as the mass-flow mean total temperature at the stage 1 nozzle trailing edge plane. Another method of determining firing temperature is defined in ISO document 2314 "Gas Turbine-Acceptance Test". The firing temperature in that case is a reference turbine inlet temperature and not generally a temperature that exists in a gas turbine cycle; it is calculated using parameters obtained in a field test. Thus, this ISO reference temperature is always less than the true firing temperature as defined by GE, above.
With reference to FIG. 2, the definition of firing temperature is illustrated by way of example. Plane A is the turbine inlet temperature, which is the average gas temperature in plane A (T.sub.A). Plane B is the firing temperature as defined by GE, and by us, which is the average gas temperature in plane B, the stage 1 nozzle trailing edge plane. Plane C identifies the ISO firing temperature which is the calculated temperature in plane C, defined by T.sub.c =f (M.sub.a, M.sub.f).
A Brayton cycle may be evaluated using such parameters as pressure, temperature, specific heat, efficiency factors, and the adiabatic compression exponent. If such an analysis is applied to a Brayton cycle, the results can be displayed as a plot a cycle efficiency versus specific output of the cycle. Output per pound of airflow is an important determination since the higher this value, the smaller the gas turbine required for the same output power. Thermal efficiency is important because it directly affects the operating fuel costs.
Many factors affect gas turbine performance. Air temperature, for example, is an important factor in gas turbine performance. Since the gas turbine receives ambient air as inlet air, its performance will be changed by anything that affects the mass flow of the air intake to the compressor; that is changes from the reference conditions of 59.degree. F. and 14.696 psia. Each turbine model has its own temperature-effect curve as it depends on the cycle parameters and component efficiencies as well as air mass flow.
It is also well known that elevated firing temperature in the gas turbine is a key element in providing higher output per unit mass flow, enabling increased combined cycle efficiency, and that for a given firing temperature, there is an optimal cycle pressure ratio which maximizes combined cycle efficiency. The optimal cycle pressure ratio trends higher with increasing firing temperature. Compressors for these turbines are thus subjected to demands for higher levels of pressure ratio, with the simultaneous goals of minimal parts count, operational simplicity, and low overall cost. Moreover, the compressor must enable this heightened level of cycle pressure ratio at a compression efficiency that augments the overall cycle efficiency. Finally, the compressor must perform in an aerodynamically and aeromechanically stable manner under a wide range of mass flow rates associated with varying power output characteristics of the combined cycle operation.
The maximum pressure ratio that the compressor can deliver in continuous duty is commonly defined in terms of a margin from the surge pressure ratio line. Compressor surge is the low frequency oscillation of flow where the flow separates from the blades and reverses flow direction through the machine, i.e., it serves as a physical limit to compressor operation at a given speed.
Conventional industrial gas turbine operational strategy has been to size the first-stage turbine nozzle throat area such that the compressor surge margin minimum limit is encountered under Power-Turn-Down, Cold-Day conditions. Nominal-flow Cold-Day may in some cases be the limiting operational condition. FIG. 3 illustrates the rationale for this prior art strategy. At the higher corrected speed associated with cold ambient conditions, the speed lines, e.g., 105%, 110%, 115% corrected speed, become closely spaced as the front stages of the compressor begin to aerodynamically choke. This causes the operating line pressure ratio to increase rapidly toward the surge line with increased speeds above 100% corrected speed. Consequently, the compressor surge margin decreases rapidly with increasing corrected speed. Accordingly, the first-stage turbine nozzle throat area has traditionally been sized such that the minimum level of surge margin experienced throughout the operational range occurs at Cold-Day conditions. This minimal level of surge pressure margin is intended to accommodate departures from new-and-clean conditions of the compressor blading, machine-to-machine variation, etc. As a result of this conventional approach, however, the surge margin at ISO-Day conditions is well in excess of the minimum safe margin, such that the rated pressure ratio delivered is well below the Cold-Day value and well below that which the compressor is capable of delivering.